For purposes of clarity, the term “conventional engine” as used in the present application refers to an internal combustion engine wherein all four strokes of the well known Otto cycle (the intake, compression, expansion and exhaust strokes) are contained in each piston/cylinder combination of the engine. Each stroke requires one half revolution of the crankshaft (180 degrees crank angle (CA)), and two full revolutions of the crankshaft (720 degrees CA) are required to complete the entire Otto cycle in each cylinder of a conventional engine.
Also, for purposes of clarity, the following definition is offered for the term “split-cycle engine” as may be applied to engines disclosed in the prior art and as referred to in the present application.
A split-cycle engine comprises:
a crankshaft rotatable about a crankshaft axis;
a compression piston slidably received within a compression cylinder and operatively connected to the crankshaft such that the compression piston reciprocates through an intake stroke and a compression stroke during a single rotation of the crankshaft;
an expansion (power) piston slidably received within an expansion cylinder and operatively connected to the crankshaft such that the expansion piston reciprocates through an expansion stroke and an exhaust stroke during a single rotation of the crankshaft; and
a crossover passage interconnecting the compression and expansion cylinders, the crossover passage including a crossover compression (XovrC) valve and a crossover expansion (XovrE) valve defining a pressure chamber therebetween.
U.S. Pat. No. 6,543,225 granted Apr. 8, 2003 to Carmelo J. Scuderi (the Scuderi patent) and U.S. Pat. No. 6,952,923 granted Oct. 11, 2005 to David P. Branyon et al. (the Branyon patent) each contain an extensive discussion of split-cycle and similar type engines. In addition the Scuderi and Branyon patents disclose details of prior versions of engines of which the present invention comprises a further development. Both the Scuderi patent and the Branyon patent are incorporated herein by reference in their entirety.
Referring to FIG. 1, a prior art split-cycle engine of the type similar to those described in the Branyon and Scuderi patents is shown generally by numeral 10. The split-cycle engine 10 replaces two adjacent cylinders of a conventional engine with a combination of one compression cylinder 12 and one expansion cylinder 14. The four strokes of the Otto cycle are “split” over the two cylinders 12 and 14 such that the compression cylinder 12 contains the intake and compression strokes and the expansion cylinder 14 contains the expansion and exhaust strokes. The Otto cycle is therefore completed in these two cylinders 12, 14 once per crankshaft 16 revolution (360 degrees CA).
During the intake stroke, intake air is drawn into the compression cylinder 12 through an inwardly opening (opening inward into the cylinder) poppet intake valve 18. During the compression stroke, compression piston 20 pressurizes the air charge and drives the air charge through the crossover passage 22, which acts as the intake passage for the expansion cylinder 14.
Due to very high volumetric compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the compression cylinder 12, an outwardly opening (opening outward away from the cylinder) poppet crossover compression (XovrC) valve 24 at the crossover passage inlet is used to control flow from the compression cylinder 12 into the crossover passage 22. Due to very high volumetric compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the expansion cylinder 14, an outwardly opening poppet crossover expansion (XovrE) valve 26 at the outlet of the crossover passage 22 controls flow from the crossover passage 22 into the expansion cylinder 14. The actuation rates and phasing of the XovrC and XovrE valves 24, 26 are timed to maintain pressure in the crossover passage 22 at a high minimum pressure (typically 20 bar or higher) during all four strokes of the Otto cycle.
A fuel injector 28 injects fuel into the pressurized air at the exit end of the crossover passage 22 in correspondence with the XovrE valve 26 opening. The fuel-air charge fully enters the expansion cylinder 14 shortly after expansion piston 30 reaches its top dead center position. As piston 30 begins its descent from its top dead center position, and while the XovrE valve 26 is still open, spark plug 32 is fired to initiate combustion (typically between 10 to 20 degrees CA after top dead center of the expansion piston 30). The XovrE valve 26 is then closed before the resulting combustion event can enter the crossover passage 22. The combustion event drives the expansion piston 30 downward in a power stroke. Exhaust gases are pumped out of the expansion cylinder 14 through inwardly opening poppet exhaust valve 34 during the exhaust stroke.
With the split-cycle engine concept, the geometric engine parameters (i.e., bore, stroke, connecting rod length, compression ratio, etc.) of the compression and expansion cylinders are generally independent from one another. For example, the crank throws 36, 38 for the compression cylinder 12 and expansion cylinder 14 respectively may have different radii and may be phased apart from one another with top dead center (TDC) of the expansion piston 30 occurring prior to TDC of the compression piston 20. This independence enables the split-cycle engine to potentially achieve higher efficiency levels and greater torques than typical four stroke engines.
The actuation mechanisms (not shown) for crossover valves 24, 26 may be cam driven or camless. In general, a cam driven mechanism includes a camshaft mechanically linked to the crankshaft. A cam is mounted to the camshaft, and has a contoured surface that controls the valve lift profile of the valve opening event [i.e., the event that occurs during a valve actuation]. A cam driven actuation mechanism is efficient, fast and may be part of a variable valve actuation system, but generally has limited flexibility.
For purposes herein a valve opening event is defined as the valve lift from its initial opening off of its valve seat to its closing back onto its valve seat versus rotation of the crankshaft during which the valve lift occurs. Also for purposes herein the valve opening event rate [i.e., the valve actuation rate] is the duration in time required for the valve opening event to occur within a given engine cycle. It is important to note that a valve opening event is generally only a fraction of the total duration of an engine operating cycle, e.g., 720 CA degrees for a conventional engine cycle and 360 CA degrees for a split-cycle engine.
Also in general, camless actuation systems are known, and include systems that have one or more combinations of mechanical, hydraulic, pneumatic, and/or electrical components or the like. Camless systems allow for greater flexibility during operation, including, but not limited to, the ability to change the valve lift height and duration and/or deactivate the valve at selective times.
Referring to FIG. 2, an exemplary prior art valve lift profile 40 for a crossover valve in a split-cycle engine is shown. Valve lift profile 40 can potentially be applied to either or both of crossover valves 24, 26 in FIG. 1. Valves 24 and 26 will be referred to below as having the same valve lift profile 40 merely for purposes of discussion.
Regardless of whether valves 24 and 26 are cam driven or actuated with a camless system, the valve lift profile 40 needs to be controlled to avoid damaging impacts when the valves 24, 26 are approaching their closed positions against their valve seats. Accordingly, a portion of the profile 40—referred to herein as the “landing” ramp 42—may be controlled to rapidly decelerate the velocity of the valves 24, 26 as they approach their valve seats. The valve lift at the start of maximum deceleration (on the descending side of the profile 40) is defined herein as the landing ramp height 44. The landing ramp duration 46 is defined herein as the duration of time from the start of the maximum deceleration of the moving valve to the point of landing on the valve seat. The velocity of the valve 24 or 26 when the valve contacts the valve seat is referred to herein as the seating velocity. For purposes herein, the “takeoff” ramp 45 is not as critical as the landing ramp 42, and can be set to any value that adequately achieves the maximum lift 48.
In cam-driven actuation systems, the landing ramp is generated by the profile of the cam. Accordingly, the landing ramp's duration in time is proportional to the engine speed, while its duration relative to crankshaft rotation (i.e., degrees CA) is generally fixed. In camless actuation systems, in general, the landing ramp is actively controlled by a valve seating control device or system.
For split-cycle engines which ignite their charge after the expansion piston reaches its top dead center position (such as in the Scuderi and Branyon patents), the dynamic actuation of the crossover valves 24, 26 is very demanding. This is because the crossover valves 24 and 26 of engine 10 must achieve sufficient lift to fully transfer the fuel-air charge in a very short period of crankshaft rotation (generally in a range of about 30 to 60 degrees CA) relative to that of a conventional engine, which normally actuates the valves for a period of at least 180 degrees CA. This means that the crossover valves 24, 26 must actuate about four to six times faster than the valves of a conventional engine.
As a consequence of the faster actuation requirements, the XovrC and XovrE valves 24, 26 of the split-cycle engine 10 have a severely restricted maximum lift (48 in FIG. 2) compared to that of valves in a conventional engine. Typically the maximum lift 48 of these crossover valves 24, 26 are in the order of 2 to 3 millimeters, as compared to about 10-12 mm for valves in a conventional engine. Consequently, both the height 44 and duration 46 of the landing ramp 42 for the XovrC and XovrE valves 24, 26, need to be minimized to account for the shortened maximum lift and faster actuation rates.
Problematically, the heights 44 of the landing ramps 42 of crossover valves 24 and 26 are so restricted that unavoidable variations in parameters that control ramp height, which are normally less significant in their effect on the larger lift profiles of conventional engines, now become critical. These parameter variations may include, but are not limited to:                1) dimensional changes due to thermal expansion of the metal valve stem and other metallic components in the valve's actuation mechanism as engine operational temperatures vary;        2) the normal wear of the valve and valve seat during the operational life of the valve;        3) manufacturing and assembly tolerances; and        4) variations in the compressibility (and resulting deflection) of hydraulic fluids (e.g. oil) in any components of the valvetrain (mainly caused by aeration).        
Referring to FIG. 3, an exemplary embodiment of a conventional cam-driven valve train 50 for a conventional engine is illustrated. For purposes herein, a valve train of an internal combustion engine is defined as a system of valve train elements, which is used to control the actuation of the valves. The valve train elements generally comprise a combination of actuating elements and their associated support elements. Also for purposes herein, the primary motion of any valve train element is defined as that motion which the element would substantially experience when the elements of the valve train are idealized to have an infinite stiffness. The actuating elements (e.g., cams, tappets, springs, rocker arms, valves and the like) are used to directly impart the primary actuation motion to the valves (i.e., to actuate the valves) of the engine during each valve opening event of the valves. Accordingly, the primary motion of the individual actuating elements in a valve train must operate at the substantially same actuation rates as the valve opening events of the valves that the actuating elements actuate. The support elements (e.g., shafts, pedestals or the like) are used to securely mount and guide the actuating elements to the engine and generally have no primary motion, although they affect the overall stiffness of the valve train system. However, the primary motion, if any, of the support elements in a valve train operate at slower rates than the valve opening events of the valves.
It should be noted that support elements may be subject to some high frequency vibration primarily caused by the high frequency movements of the actuating elements of a valve train, which apply forces to the support elements during operation. The high frequency vibrations are a consequence of the actuating and support elements of the valve train having a finite stiffness, and are not part of the primary motion. However, the displacement induced by this vibration alone will have a magnitude that is substantially less than the magnitude of the primary motion of the actuating elements in the valve train, typically by an order of magnitude or less.
Valve train 50 actuates an inwardly opening poppet valve 52 having a valve head 54 and a valve stem 56. Located at the distal end of the valve stem 56 is the valve tip 58, which abuts against a tappet 60. Spring 62 holds the valve head 54 securely against a valve seat 64 when the valve 52 is in its closed position. Cam 66 rotates to act against the tappet 60 in order to depress spring 62 and lift the valve head 54 off of its valve seat 64. In this exemplary embodiment, valve 52, spring 62, tappet 60 and cam 66 are actuating elements. Though no associated support elements are illustrated, one skilled in the art would recognize that they would be required. Cam 66 includes a cylindrical portion, generally referred to as the base circle 68, which does not impart any linear motion to the valve 52. Cam 66 also includes a lift (or eccentric) portion 70 that imparts the linear motion to the valve 52. The contour of the cam's eccentric portion 70 controls the lift profile of valve 52. The effects of the aforementioned dimensional changes due to thermal expansion are compensated for by including a preset clearance space (or clearance) 72.
For purposes herein, the terms “valve lash” or “lash”: are defined as the total clearance existing within a valve train when the valve is fully seated. The valve lash is equal to the total contribution of all the individual clearances between all individual valve train elements (i.e., actuating elements and support elements) of a valve train
In this particular embodiment, the clearance 72 is the distance between the base circle 68 of cam 66 and the tappet 60. Also note that, in this particular embodiment, the clearance 72 is substantially equal to the valve lash of the valve train, i.e., the total contribution of all the clearances that exist between the valve's distal tip 58, when the valve 52 is fully seated on the valve seat 64, and the cam 66.
To compensate for the thermal effects on the inwardly opening valve 52, the clearance 72 is set at its maximum tolerance when the engine is cold. When the engine heats up, the valve's stem 56 will expand in length and reduce the clearance 72, but will not abut against the cam's base circle 68 (i.e., will not reduce the clearance 72 to zero). Accordingly, as the clearance 72 is reduced, valve 52 is extended further into the cylinder (not shown) when the valve 52 is open. Note however that, even as the clearance 72 is reduced, valve 52 remains seated against its valve seat when the valve 52 is closed.
However, as mentioned above, crossover valves, such as valves 24, 26 in split-cycle engine 10, have lift profiles that include much smaller landing ramp heights compared to that of a conventional engine. This would be true whether the valves were inwardly opening or outwardly opening, so long as the duration of valve actuation [i.e., the valve opening event] was short relative to that of a valve on a conventional engine, for example, a valve with a duration of actuation of approximately 3 ms and 180 degrees of crank angle, or less. In the case of such fast actuating, cam driven, inwardly opening valves, the valve's distal tip must engage the cam's landing ramps in order to have a controlled landing and safe seating velocity, and any fixed valve lash for such inwardly opening crossover valves must necessarily be set proportionally small. Problematically, variations in a set valve lash due to thermal expansion effects may actually be greater than the ramp height required for such valves. This means that if the valve lash is set large enough to account for thermal expansion, the tips of these inwardly opening crossover valves could miss the landing ramp altogether, which would cause the valves to repeatedly crash against their valve seats and prematurely damage the valves. Additionally, if the valve lash is set small enough to guarantee engagement with the landing ramp at all operating temperatures, the tips of the valves could expand enough to abut against the base circle of the cam, which would force the inwardly opening crossover valves open even when the valves should be in their closed position.
Moreover, the large lash setting would generate a shorter valve lift duration and the small lash setting would generate a lengthened valve lift duration. In either case, the range of variation of the valve opening event can be larger than desirable. It is desirable to contain the range of the valve opening event to a manageable level.
Referring to FIG. 4, an exemplary embodiment of a conventional engine cam driven valve train 73 having an automatically adjustable valve lash is illustrated. The valve train 73 actuates inwardly opening poppet valve 74. The valve train 73 includes cam 76, pivoting lever arm 78 and spring 80 as valve train actuating elements which actuate valve 74 during each cycle. The effects of thermal expansion and other parameters mentioned above are addressed by adding a lash adjuster assembly. For the lash adjuster assembly, an active lash control device, such as a hydraulic lash adjuster (HLA) 82 has been used. The hydraulic lash adjuster (HLA) 82 also functions as a support element associated with lever arm 78. As is known in the art, as valve lash in the valve train varies, HLA 82 hydraulically adjusts the position of lever arm 78 to compensate and bring the valve lash to zero (in this particular embodiment, the valve lash would be any clearance between the cam 76 and the lever arm 78, as well as any clearance between the lever arm 78 and the distal tip of the stem of valve 74).
Because lever arm 78 is one of the valve train 73 actuating elements (i.e., is an element that directly actuates the inwardly opening valve 74 during each cycle and is used to directly impart the primary actuation motion to the valve 74), there is an unavoidable tradeoff between the lever arm's minimum mass required for adequate stiffness (ratio of force applied to a point on the lever arm to the deflection of that point caused by that force) and the maximum mass allowable for high speed operation. That is, if the mass of lever arm 78 is too small, it will not be able to actuate valve 74 without undue bending and/or deformation. Additionally, if the mass of lever arm 78 is too large, it will be too heavy to actuate valve 74 at its maximum operating speed. For any particular valve train actuating element, if the minimum mass required for adequate stiffness exceeds the maximum mass allowable for maximum operating speed, the element cannot be used in the valve train. Generally, in a conventional engine, the requirements for stiffness and speed are not so demanding as to preclude the use of lever arm 78 in valve train 73.
However, as mentioned above, crossover valves 24, 26 must actuate about four to six times faster than the valves of a conventional engine, which means the actuating elements of the valve train system must operate at extremely high and rapidly changing acceleration levels relative to that of a conventional engine. These operating conditions would severely restrict the maximum mass of lever arm 78 in valve train 73.
Additionally, crossover valves 24, 26 must open against very high pressures in the crossover passage 22 compared to that of a conventional engine (e.g., 20 bar or higher), which exacerbates the stiffness requirements on the valve train system. Also, bending is a problem on elements such as lever arm 78 because the actuation force in one direction is concentrated in the median section of the element (i.e., where cam 76 engages lever arm 78) and all opposing reactionary forces are concentrated at the end sections of the lever arm (i.e., where HLA 82 and the tip of valve 74 engage opposing ends of lever arm 78). Moreover, this bending problem would increase proportionally as the length of the lever arm 78 increases. Accordingly, if the engine illustrated in prior art FIG. 4 were subjected to the higher pressures and severe actuation rates encountered in split-cycle engine 10, the stiffness and mass of lever arm 78 in valve train 73 would have to be substantially increased, therefore restricting the overall actuation rate of valve train 73.
Generally too, prior art HLAs (such as HLA 82), because of the compressibility of oil contained therein, are normally one of the main contributing factors in reducing valve train stiffness which, in turn, limits the maximum engine operating speed at which the valve train can safely operate. Therefore, a prior art HLA 82 connected to a lever arm 78, as shown in valve train 73, cannot be implemented with the split cycle engine 10, in which the valves need to actuate much more rapidly, and the HLA 82 must be much stiffer than those in a conventional engine.
There is a need therefore, for a valve lash adjustment system for cam driven valves of a split-cycle engine, which can both (a) handle the high speed and stiffness requirements necessary to safely actuate the valves; and (b) automatically compensate for such unavoidable factors as thermal expansion of actuation components, valve wear, and manufacturing tolerances that cause variations in the lash.